US4140044A - Long stroke, large bore, low friction hydraulic actuators - Google Patents

Long stroke, large bore, low friction hydraulic actuators Download PDF

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Publication number
US4140044A
US4140044A US05/710,767 US71076776A US4140044A US 4140044 A US4140044 A US 4140044A US 71076776 A US71076776 A US 71076776A US 4140044 A US4140044 A US 4140044A
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United States
Prior art keywords
piston
piston rod
undercuts
fluid
guide member
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Expired - Lifetime
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US05/710,767
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English (en)
Inventor
Joseph Biller
Willard D. Kaiser
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Raytheon Co
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Singer Co
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Priority to US05/710,767 priority Critical patent/US4140044A/en
Priority to GB19623/77A priority patent/GB1572956A/en
Priority to CA278,257A priority patent/CA1054487A/en
Priority to JP6391777A priority patent/JPS5316172A/ja
Priority to DE19772725434 priority patent/DE2725434A1/de
Application granted granted Critical
Publication of US4140044A publication Critical patent/US4140044A/en
Assigned to LINK FLIGHT SIMULATION CORPORATION, KIRKWOOD INDUSTRIAL PARK, BINGHAMTON, NY 13902-1237, A DE CORP. reassignment LINK FLIGHT SIMULATION CORPORATION, KIRKWOOD INDUSTRIAL PARK, BINGHAMTON, NY 13902-1237, A DE CORP. ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: SINGER COMPANY, THE, A NJ CORP.
Assigned to CAE-LINK CORPORATION, A CORP. OF DE. reassignment CAE-LINK CORPORATION, A CORP. OF DE. MERGER (SEE DOCUMENT FOR DETAILS). DECEMBER 1, 1988, DELAWARE Assignors: CAE-LIN CORPORATION, A CORP. OF DE (CHANGED TO), LINK FACTICAL MILITARY SIMULATION CORPORATION, A CORP. OF DE, LINK FLIGHT SIMULATION CORPORATION, A DE CORP., LINK TRAINING SERVICES CORPORATION, A CORP. OF DE (MERGED INTO)
Assigned to HUGHES AIRCRAFT COMPANY reassignment HUGHES AIRCRAFT COMPANY ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: CAE-LINK CORPORATION
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J10/00—Engine or like cylinders; Features of hollow, e.g. cylindrical, bodies in general
    • F16J10/02—Cylinders designed to receive moving pistons or plungers

Definitions

  • the invention relates to hydraulic cylinder devices.
  • the invention concerns long stroke, large bore, low friction hydraulic actuators. (Herein, “Friction” refers to "turn-around friction”.)
  • Hydraulic actuators having cylinder bores approaching two inches, and piston strokes of twelve inches are fairly standard devices. (Piston stroke may be defined as the maximum excursion of the piston rod from the cylinder). In motion system simulators using hydraulic actuators, it is not uncommon to find hydraulic cylinders having bores in excess of two inches and stroke lengths as great as sixty inches or more. In general it may be said that cylinders having bores in excess of two inches and stroke lengths of eighteen inches or greater may be defined as large bore, long stroke hydraulic actuators.
  • Deceleration may be controlled with internal cushioning.
  • Cushioning can be provided, and usually is, at both ends of the cylinder. Cushioning at one end slows the piston as it approaches full retraction and at the other end it slows the piston as it nears full extension.
  • the typical internal cushion comprises a spear and cavity arrangement. When the piston moves, hydraulic fluid is forced out in front of its motion and exits through the cavity. As the piston approaches the end of its travel, the spear enters and begins to block part of the cavity opening. The blocking effect continues as the spear penetrates further into the cushion cavity. As the blockage increases, less fluid flows from in front of the piston into and through the cavity. This exerts a cushioning force on the piston, gradually slowing it, and stopping it when the spear completely fills the cavity.
  • the invention has as its objective the production of a large bore, long stroke, low friction hydraulic actuator having friction levels of one hundred pounds or less and utilizing internal deceleration cushioning which is not subject to wear and degradation during operation.
  • Narrow, low friction, piston rings permit the use of large piston-to-cylinder clearance, thus avoiding interference fits resulting from manufacturing tolerance build-up, without objectionably increasing oil leakage flow or friction levels. It may be noted here that the friction caused by these piston rings is independent of the number of ring elements used. Since each piston ring will produce a pressure drop, fluid leakage can be minimized, without excessively increasing friction, by using several sealing rings.
  • Piccolo hole cushioning devices are utilized as a safety feature to prevent injury to personnel and equipment should the hydraulic system fail and produce catastrophic deceleration as the piston reaches one or the other ends of its excursion.
  • a technique for the incorporation of a permanent magnet and a magnetostrictive element within the hydraulic cylinder itself is disclosed so that positional information as to the location of the piston along its stroke path may be accurately known at all times.
  • FIG. 1 shows the reactive forces at the rod seal and piston associated with the fully extended, horizontal, hydraulic actuator.
  • FIG. 2 is an illustration of the result of elastic deformation of the piston rod on the rod end-seal bearings.
  • FIGS. 3A and 3B are sketches of the dual hydrostatic bearing.
  • FIGS. 4A and 4B illustrate the use of multiple, narrow sealing rings and the pressure distribution across them.
  • FIG. 5 is a sketch of the long stroke, large bore, low friction hydraulic actuator.
  • Piston rod end-seals are often themselves another source of significant friction force. These friction forces become even more significant when the reactive forces noted above (that is, those forces resulting when the piston rod is fully extended, the cylinder horizontal, and rod and cylinder are supported at their extreme ends), are accompanied by elastic deformation of the piston rod associated with the large bending moment produced by these forces. Elastic deflections of the piston rod tend to limit the available length of the rod end-seals as will be shown herein.
  • friction forces F are developed between the piston rod and the cylinder body due to forces resulting both from the weight of the cylinder and from the normal seal forces due to a pressure drop across the seals. Both forces can be significant.
  • the reaction force at the piston head would be about 430 pounds.
  • the reaction force at the rod bearing would be about 440 pounds. Break-away friction force due only to the weight of the cylinder would thus be about 130 pounds.
  • seal friction levels in a pressurized cylinder exceed 200 pounds and may easily approach 1000 pounds when made in production quantities.
  • the method conceived for eliminating the angular deflection problem was to use dual hydrostatic bearings to support the piston rod. See FIG. 3. With two bearings supporting the rod, a floating piston design was possible leading to drastic reduction in piston bearing friction forces at the cylinder wall. Analysis of the elastic deflection of the piston rod shows that the angular deflection of the rod at the seal where the rod exits from the cylinder is greater than the angular deflection experienced by the rod as it passes through the seal inboard of the cylinder. It may be shown that the maximum seal or bearing length that can be used without mechanical interference is equal to the angular misalignment (in radians) times the bearing clearance.
  • the second bearing provides additional sealing area.
  • the length of this additional sealing area can be greater than the length of the sealing area of the first bearing at the cylinder head end because of the reduced angular deflection of the piston rod as it passes through this inboard bearing.
  • Hydrostatic bearings provide sealing by use of close mechanical tolerances in conjunction with the hydraulic fluid's viscosity. This type of sealing may be termed viscous or dynamic sealing.
  • the leakage rate of the fluid may be shown to be proportional to the clearance cubed, and inversely proportional to the length of the seal. Leakage is also affected by the centering of the rod within the bearing, bearing lower by a factor of as much as 2.5 when the rod is centered in the bearing as compared to the case when the rod is fully eccentric to the bearing.
  • the maximum seal length was increased by a factor of 3 which thus decreased the leakage rate by the same factor due to the increase in the effective sealing length. Since either hydrostatic bearing also functions to center the shaft within the bearing, decreased leakage by another factor of up to 2.5 was experienced. Thus, total leakage decreased by a factor of approximately 7.5 as compared to a conventional single bearing. A design goal leakage rate of one gallon per minute with a source pressure of 1000 psi was easily achieved. Actual levels of leakage were measured at 0.4-0.6 gallons per minute.
  • One or many narrow sealing rings may be used to reduce contact force. If several identical rings are used as shown in FIG. 4, the pressure drop across each ring P R will be total pressure drop P 1 - P 2 divided by the number of rings N. This relationship may be further expressed in the folowing equation for net radial force F R ##EQU1## where ⁇ d is the length of the seal along the hydraulic cylinder circumference.
  • W is the width of a sealing ring
  • P 1 - P 2 is the total pressure drop across the sealing area
  • the total radial force is N times the force of a single ring or
  • FIG. 4A indicates the use of multiple sealing rings on a hydraulic actuator piston to provide sealing between the piston head 12 and the cylinder wall 16 of the actuator.
  • FIG. 4B illustrates the pressure distribution across the multiple sealing rings. The lengths of the arrows are indicative of the drop in average contact pressure applied to each ring.
  • the average contact pressure P n at any n th ring is determined by the equation: ##EQU2## Since the average contact force at any ring is equal to P n over the sealing area of the ring, the contact force falls off in the same manner as the contact pressure.
  • narrow sealing rings was evaluated using commercially available, cast-iron, automotive piston rings having a width of 0.078 inches. With nominal head end and cap end pressures of near 500 psi, and a differential pressure of 120 psi, the static friction force of a single ring was measured at 11 pounds. Leakage was measured at 0.033 gpm, using 190 ssu oil (100° F.) operating at about 140° F.
  • the cushion is normally formed by some form of a spear and cavity arrangement.
  • the spear enters the cavity as the piston head is exercised near the end of its excursion, gradually blocking the oil and forming the cushion.
  • Close mechanical tolerances are required to form an effective seal, especially at the rod side of the piston head.
  • the cushioning technique most suitble for simulator applications and the like was one in which, as the piston reaches the end of its excursion, oil is forced out of the cylinder through small holes in the cylinder wall. With this design, the piston head progressively blocks off the holes as it travels toward the end of the cylinder. By proper sizing and positioning of the holes, accurate control of the cushioning action can be achieved.
  • a cylinder with an array of cushioning holes has been sometimes referred to as, "a metering orifice cylinder”, or more simply and more frequently it has been denoted, "a piccolo cushion cylinder".
  • Piccolo cushioning techniques are conventionally applied to shock absorber and door check valve cylinders, railroad couplers and the like.
  • Piston displacement is measured in terms of time elapsed between the initiation of the circumferential magnetic field and the detection of the torsion pulse. Since the torsion pulse is established by the location of the annular magnet along the sonic delay line, the time will vary as the piston is exercised within the hydraulic cylinder varying the position at the magnet along the length of the line.
  • the sonic wave guide displacement transducer is illustrated in FIG. 5 and will be further discussed in light of that figure.
  • FIG. 5 illustrates an embodiment of a hydraulic actuator as comprising a tubular member 50 defining a fluid chamber 51.
  • the tubular member 50 is closed at one end (the cap end) by a closure member 52.
  • Closure member 52 is secured to the end of tubular member 50 in a flid-tight manner using means well known to the art.
  • a suitable mounting fixture, not shown, is formed at the end of closure member 52.
  • Fluid chamber 51 Within fluid chamber 51 is located a piston comprising piston head 53 and an operating piston rod 54.
  • Fluid chamber 57 may be formed by an undercut in the inner wall of sleeve 80.
  • Sleeve 80 is secured to tubular member 50 in a fluid-tight manner.
  • openings 56 provide a fluid flow path to a third fluid chamber 58, also coaxial to fluid chamber 51.
  • Fluid chmber 58 may be formed by an undercut in the inner wal of sleeve 81.
  • Sleeve 81 is secured to tubular member 50 in a fluid-tight manner.
  • a guide member 59 is inserted into tubular member 50. The end of tubular member 50 and the flange end of guide member 59 are joined in a fluid tight manner.
  • Guide member 59 is provided with an axially extending opening 60 which axially, slidably receives piston rod member 54.
  • a fluid port 61 is carried by guide member 59 and is connected to axial opening 60 by means of restricted orifice 62. Fluid port 61 is adapted to be connected to a source of high pressure fluid flow not shown. Restricted orifice 62 communicates the high pressure fluid to axial opening 60 where, in conjunction with undercut pockets 65 (FIG. 3), a hydrostatic bearing is formed. Becauseorifice 62 is restricted, fluid flow is limited but high pressure is maintained in the bearing areas.
  • Fluid port 61 communicates by way of fluid duct 63 with a fourth fluid chamber 64, which may also be coaxial with fluid chamber 51.
  • a fluid flow path from fluid chamber 64 to axial opening 60 is provided by restricted orifice 66.
  • undercut pockets 67 FIG. 3
  • a second hydrostatic bearing is formed as the high pressure fluid exits from restricted orifice 66.
  • Guide member 59 is provided with an undercut 68, thereby enlarging axial opening 60 in the region of the undercut.
  • Undercut 68 permits the elastic deformation of piston rod 54 without contact being made to guide member 59.
  • Undercut 68 also obviates the need for maintaining close manufacturing tolerances along the full length of axial opening 60.
  • Undercut 68 is centrally located between hydrostatic bearing pockets 65 and 67.
  • bearing pockets 65 and 67 there is essentially no fluid flow between bearing pockets 65 and 67 since the fluid pressure at one set of bearing pockets equally opposes the tendency of fluid to flow from the opposite set of pockets. Further, fluid flow from bearing pockets 67 toward the cap end of fluid chamber 51 is opposed by the fluid pressure within chamber 51. Such flow as there is, is relatively insignificant and is directed via gland duct 82 through fluid port 69 to a scavening means not shown.
  • a low friction wiper seal may be utilized at the head end of the cylinder to the left of gland duct 82 illustrated in FIG. 5 to provide a wiping action on the surface of piston rod 54.
  • An elongated chamber 70 is provided in piston rod 54 and piston head 53. Chamber 70 is coaxial to both piston head 53 and piston rod 54. Chamber 70 axially slidably receives an elongated, magnetostrictive, sonic delay line 71.
  • An annular magnet 72 slidebly encompasses sonic delay line 71 and is affixed to piston head 53 in such a manner that activation of the hydraulic actuator causes magnet 72 to move slideably along the length of snic delay line 71 as the piston 53 is exercised.
  • Sonic delay line 71 is affixed at one end to transducer 73.
  • Transducer 73 is mounted internal to closure member 52 and is thus not exposed to the environment surrounding the hydraulic actuator. Transducer 73 is provided with an electrical connector 74 to provide external electrical access to transducer 73.
  • Fluid chamber 51 is provided with fluid ports 75 and 76. Fluid ports 75 and 76 are adapted to be connected to a source of fluid flow through a conventional directional change means not shown. Fluid chamber 57 is provided with fluid port 77. Fluid chamber 58 is provided with fluid port 78. Fluid ports 77 and 78 are adapted to be connected through a conventional directional change means to return fluid to the source of fluid flow which is not shown herein.
  • piston head 53 will move axially to the left as the hydraulic actuator is shown in FIG. 5. This movement of piston head 53 will produce a corresponding axial movement of the rod member 54 since these members are all secured together.
  • piston head 53 and piston rod 54 Upon fluid being directed through fluid port 75 into the head-end of fluid chamber 51, piston head 53 and piston rod 54 will be moved to the right toward the position illustrated in FIG. 5. In the course of this latter motion, fluid will be exhausted from the cap end of fluid chamber 51 through openings 55 into chamber 57 and fluid port 77.
  • piston head 53 does not make slideable contact with the walls of fluid chamber 51, pressure difference between the head end and the cap end of fluid chamber 51 is maintained by the provision of narrow piston rings 79 located circumferentially about piston head 53.
  • fluid be exhausted through openings 56 or 55 provides a safe, controlled, cushioned stop at either end of the piston travel.
  • fluid is exhausted through plurality of openings 56.
  • ring 79 passes the largest of the openings 56. This action restricts the flow of the fluid being exhausted to that which is possible through the remaining smaller openings 56 and the piston is slowed.
  • ring 79 continues past each successive hole, further restricting the flow of fluid, further slowing the travel of the piston until finally all flow is blocked when the last hole is passed.
  • the remaining fluid acts as a cushion to safely and controllably stop the piston travel.
  • piston 53 advances to the right of the illustration in FIG. 5 and the fluid is exhausted through plurality of openings 55.
  • Hydraulic fluid from a pressure source not shown is supplied to fluid port 61.
  • the hydraulic fluid is then directed into axial opening 60 through restricted orifices 62 and via fluid duct 63, fluid chamber 64, and restricted orifices 66.
  • the restricted orifices are used to control the flow of oil into the hydrostatic bearing pockets 65 and 67.
  • commercially available Lee Jets registered trademark of the Lee Company or the equivalent may be used.
  • a second consideration is the degree of elastic deformation of piston rod 54 when it is extended and again subjected to lateral loads.
  • the length and separation of the hydrostatic bearings may be established to maintain minimal non-contacting clearance between the elastically deformed piston rod and guidemember 59 at any point along axial opening 60.
  • the hydrostatic bearing lengths and separations are also chosen such that piston head 53 may be non contacting on the walls of fluid chamber 51 and that no increase in frictional force is exerted by piston 53 via ring 79 against the walls of fluid chamber 51 when a lateral load is applied, such as may be experienced on a fully extended hydraulic actuator supported at its extreme ends.
  • the length of the bearing seals is also determined by the hydraulic fluid leakage rate which may be tolerated. Such viscous sealing has been discussed earlier.
  • Minimum clearances between the piston rod 54 and guide member 59 are determined not only by elastic deflections but also by manufacturing tolerances.
  • the embodiment of the hydraulic actuator described herein, as an example of the invention may be taken as a 56 inch stroke, 3.5 inch bore, 2.5 inch rod actuator. Calculations based on elastic deflection and constraints upon the length of stop tube 59 indicated that a dual hydrostatic bearing having a head end bearing length of 3 inches and an inboard bearing length of 6 inches could be utilized with diametral clearances as small as 0.0025 inches, if manufacturing angular alignment errors were kept small.
  • the maximum diametral clearance that can be used is determined by the leakage flow rate which may be tolerated from the head end of the actuator.
  • the hydrostatic bearing fluid flow is restricted and may be treated as negligible relative to the leakage flows of the fluid in fluid chamber 51.
  • Equation 5 gives a calculated leakage flow rate of 2.45 in. 3 /sec (0.64 gpm).
  • a prototype cylinder was constructed according to the teachings herein which consistently operated with total turn-around friction levels of 75 pounds or less. Using low viscosity hydraulic fluid (80 to 90 ssu), the combined leakage flow from the head end of the cylinder was measured to be on the order of 0.4 to 0.6 gallons gpm, significantly less than the expected 1.0 gpm.
  • a long stroke, large bore, low friction hydraulic actuator has been described.
  • very narrow seals are effective in reducing friction, a large number of such narrow seals theoretically producing much less friction than a single seal of equivalent length; that where piston rod elastic deflections are so significant that they are a major factor in determining rod end leakage, the use of a dual hydrostatic bearing configuration can reduce rod end leakage by a factor of at least 3:1; and that with dual hydrostatic bearings the piston head and cylinder wall do not have to carry forces produced by rod bending moments necessary to support the weight of the cylinders, the necessary reaction moments being produced by the dual hydrostatic bearings.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • Actuator (AREA)
US05/710,767 1976-07-30 1976-07-30 Long stroke, large bore, low friction hydraulic actuators Expired - Lifetime US4140044A (en)

Priority Applications (5)

Application Number Priority Date Filing Date Title
US05/710,767 US4140044A (en) 1976-07-30 1976-07-30 Long stroke, large bore, low friction hydraulic actuators
GB19623/77A GB1572956A (en) 1976-07-30 1977-05-10 Long stroke large bore low friction hydraulic actuators
CA278,257A CA1054487A (en) 1976-07-30 1977-05-12 Long stroke, large bore, low friction hydraulic actuators
JP6391777A JPS5316172A (en) 1976-07-30 1977-05-31 Hydraulic actuator having large diameter* long stroke* and low friction
DE19772725434 DE2725434A1 (de) 1976-07-30 1977-06-04 Hydraulischer arbeitszylinder

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US05/710,767 US4140044A (en) 1976-07-30 1976-07-30 Long stroke, large bore, low friction hydraulic actuators

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JP (1) JPS5316172A (US08088918-20120103-C00476.png)
CA (1) CA1054487A (US08088918-20120103-C00476.png)
DE (1) DE2725434A1 (US08088918-20120103-C00476.png)
GB (1) GB1572956A (US08088918-20120103-C00476.png)

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US4463563A (en) * 1978-06-19 1984-08-07 The Cessna Aircraft Company Rephasing cylinder construction
US4543649A (en) * 1983-10-17 1985-09-24 Teknar, Inc. System for ultrasonically detecting the relative position of a moveable device
US4910961A (en) * 1987-05-21 1990-03-27 Vertran Manufacturing Company Hydraulic door opening or closing device
US4944215A (en) * 1988-12-13 1990-07-31 Nimmo Frank D Fluid actuated cylinder assembly with a floating cylinder head
US5107677A (en) * 1987-05-21 1992-04-28 Vertran Manufacturing Company Hydraulic door actuator
US5161957A (en) * 1987-05-21 1992-11-10 Vertran Manufacturing Company Hydraulic door actuator
US5322025A (en) * 1992-05-29 1994-06-21 Steelcase Inc. Adjustable dual worksurface support
US5477771A (en) * 1993-08-10 1995-12-26 Black; Philip B. Hydraulic cylinder assembly
US5952823A (en) * 1996-03-22 1999-09-14 Mts Systems Corporation Magnetostrictive linear displacement transducer for a shock absorber
US6401883B1 (en) 1999-09-22 2002-06-11 Mts Systems Corporation Vehicle suspension strut having a continuous position sensor
US6729419B1 (en) 1999-05-28 2004-05-04 Smith International, Inc. Electro-mechanical drilling jar
DE102006047966A1 (de) * 2006-07-07 2008-01-10 Asm Automation Sensorik Messtechnik Gmbh Gleitelement mit Positionsgeber
US20090084257A1 (en) * 2007-09-28 2009-04-02 Caterpillar Inc. Hydraulic cylinder having multi-stage snubbing valve
US20090107249A1 (en) * 2007-10-24 2009-04-30 Thaddeus Schroeder Means and method of sensing pressure using magnetostrictive electrical condutors
US20090271998A1 (en) * 2008-05-02 2009-11-05 Carlen Controls, Inc. Linear Position Transducer With Wireless Read Head
US20150285243A1 (en) * 2014-04-07 2015-10-08 i2r Solutions USA LLC Hydraulic Pumping Assembly, System and Method
US20160169219A1 (en) * 2014-12-15 2016-06-16 Sustainable Waste Power Systems, Inc. Pumps, pump assemblies, and methods of pumping fluids
CN112360838A (zh) * 2020-10-28 2021-02-12 辽宁工程技术大学 一种高频响、高精度和低摩擦的数字流体缸

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JPS6042155B2 (ja) * 1979-12-24 1985-09-20 三井造船株式会社 クレ−ンの吊荷振れ制動装置
EP0103852B1 (de) * 1982-09-20 1986-05-28 IMI Norgren AG Linearantrieb
DE102009001770A1 (de) * 2009-03-24 2010-09-30 Zf Friedrichshafen Ag Druckmittelzylinder mit optimierter Kolbenführung

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US3311030A (en) * 1965-02-09 1967-03-28 Halstead Metal Products Inc Self-aligning packing gland arrangements
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US3742822A (en) * 1971-08-03 1973-07-03 Union Carbide Corp Close clearance viscous fluid seal system
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Cited By (21)

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Publication number Priority date Publication date Assignee Title
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Also Published As

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CA1054487A (en) 1979-05-15
JPS5618809B2 (US08088918-20120103-C00476.png) 1981-05-01
JPS5316172A (en) 1978-02-14
DE2725434A1 (de) 1978-02-02
GB1572956A (en) 1980-08-06

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